Split four stroke engine

ABSTRACT

An engine has a crankshaft rotating about a crankshaft axis of the engine. A power piston is slidably received within a first cylinder and operatively connected to the crankshaft such that the power piston reciprocates through a power stroke and an exhaust stroke of a four stroke cycle during a single rotation of the crankshaft. A compression piston is slidably received within a second cylinder and operatively connected to the crankshaft such that the compression piston reciprocates through an intake stroke and a compression stroke of the same four stroke cycle during the same rotation of the crankshaft. A gas passage interconnects the first and second cylinders. The gas passage includes an inlet valve and an outlet valve defining a pressure chamber therebetween. The outlet valve permits substantially one-way flow of compressed gas from the pressure chamber to the first cylinder. Combustion is initiated in the first cylinder between 0 degrees and 40 degrees of rotation of the crankshaft after the power piston has reached its top dead center position.

CROSS REFERENCE TO RELATED APPLICATIONS

This patent application is a continuation application of U.S.application Ser. No. 10/615,550, filed Jul. 8, 2003, entitled “SPLITFOUR STROKE ENGINE”, which is a continuation application of U.S.application Ser. No. 10/139,981 (now U.S. Pat. No. 6,609,371), filed May7, 2002, entitled “SPLIT FOUR STROKE ENGINE”, which is a continuationapplication of U.S. application Ser. No. 09/909,594 (now U.S. Pat. No.6,543,225), filed Jul. 20, 2001, entitled “SPLIT FOUR STROKE CYCLEINTERNAL COMBUSTION ENGINE”, all of which are herein incorporated byreference in their entirety.

FIELD OF THE INVENTION

The present invention relates to internal combustion engines. Morespecifically, the present invention relates to a four-stroke cycleinternal combustion engine having a pair of offset pistons in which onepiston of the pair is used for the intake and compression strokes andanother piston of the pair is used for the power and exhaust strokes,with each four stroke cycle being completed in one revolution of thecrankshaft.

BACKGROUND OF THE INVENTION

Internal combustion engines are any of a group of devices in which thereactants of combustion, e.g., oxidizer and fuel, and the products ofcombustion serve as the working fluids of the engine. The basiccomponents of an internal combustion engine are well known in the artand include the engine block, cylinder head, cylinders, pistons, valves,crankshaft and camshaft. The cylinder heads, cylinders and tops of thepistons typically form combustion chambers into which fuel and oxidizer(e.g., air) is introduced and combustion takes place. Such an enginegains its energy from the heat released during the combustion of thenon-reacted working fluids, e.g., the oxidizer-fuel mixture. Thisprocess occurs within the engine and is part of the thermodynamic cycleof the device. In all internal combustion engines, useful work isgenerated from the hot, gaseous products of combustion acting directlyon moving surfaces of the engine, such as the top or crown of a piston.Generally, reciprocating motion of the pistons is transferred to rotarymotion of a crankshaft via connecting rods.

Internal combustion (IC) engines can be categorized into spark ignition(SI) and compression ignition (CI) categories. SI engines, i.e. typicalgasoline engines, use a spark to ignite the air-fuel mixture, while theheat of compression ignites the air fuel mixture in CI engines, i.e.,typically diesel engines.

The most common internal-combustion engine is the four-stroke cycleengine, a conception whose basic design has not changed for more than100 years old. This is because of its outstanding performance as a primemover in the ground transportation industry. In a four-stroke cycleengine, power is recovered from the combustion process in four separatepiston movements (strokes) of a single piston. For purposes herein, astroke is defined as a complete movement of a piston from a top deadcenter position to a bottom dead center position or vice versa.Accordingly, a four-stroke cycle engine is defined herein to be anengine which requires four complete strokes of one or more pistons forevery power stroke, i.e. for every stroke that delivers power to acrankshaft.

Referring to FIGS. 1-4, an exemplary embodiment of a prior art fourstroke cycle internal combustion engine is shown at 10. For purposes ofcomparison, the following four FIGS. 1-4 describe what will be termed aprior art “standard engine” 10. As will be explained in greater detailhereinafter, this standard engine 10 is an SI engine with a 4 inchdiameter piston, a 4 inch stroke and an 8 to 1 compression ratio. Thecompression ratio is defined herein as the maximum volume of apredetermined mass of an air-fuel mixture before a compression stroke,divided by the volume of the mass of the air-fuel mixture at the pointof ignition. For the standard engine, the compression ratio issubstantially the ratio of the volume in cylinder 14 when piston 16 isat bottom dead center to the volume in the cylinder 14 when the piston16 is at top dead center.

The engine 10 includes an engine block 12 having the cylinder 14extending therethrough. The cylinder 14 is sized to receive thereciprocating piston 16 therein. Attached to the top of the cylinder 14is the cylinder head 18, which includes an inlet valve 20 and an outletvalve 22. The cylinder head 18 cylinder 14 and top (or crown 24) of thepiston 16 form a combustion chamber 26. On the inlet stroke (FIG. 1), afuel air mixture is introduced into the combustion chamber 26 through anintake passage 28 and the inlet valve 20, wherein the mixture is ignitedvia spark plug 30. The products of combustion are later exhaustedthrough outlet valve 22 and outlet passage 32 on the exhaust stroke(FIG. 4). A connecting rod 34 is pivotally attached at its top distalend 36 to the piston 16. A crankshaft 38 includes a mechanical offsetportion called the crankshaft throw 40, which is pivotally attached tothe bottom distal end 42 of connecting rod 34. The mechanical linkage ofthe connecting rod 34 to the piston 16 and crankshaft throw 40 serves toconvert the reciprocating motion (as indicated by arrow 44) of thepiston 16 to the rotary motion (as indicated by arrow 46) of thecrankshaft 38. The crankshaft 38, is mechanically linked (not shown) toan inlet camshaft 48 and an outlet camshaft 50, which precisely controlthe opening and closing of the inlet valve 20 and outlet valve 22respectively.

The cylinder 14 has a centerline (piston-cylinder axis) 52, which isalso the centerline of reciprocation of the piston 16. The crankshaft 38has a center of rotation (crankshaft axis) 54. For purposes of thisspecification, the direction of rotation 46 of the crankshaft 38 will bein the clockwise direction as viewed by the reader into the plane of thepaper. The centerline 52 of the cylinder 14 passes directly through thecenter of rotation 54 of the crankshaft 38.

Referring to FIG. 1, with the inlet valve 20 open, the piston 16 firstdescends (as indicated by the direction of arrow 44) on the intakestroke. A predetermined mass of an explosive mixture of fuel (gasolinevapor) and air is drawn into the combustion chamber 26 by the partialvacuum thus created. The piston continues to descend until it reachesits bottom dead center (BDC), the point at which the piston is farthestfrom the cylinder head 18.

Referring to FIG. 2, with both the inlet 20 and outlet 22 valves closed,the mixture is compressed as the piston 16 ascends (as indicated by thedirection of arrow 44) on the compression stroke. As the end of thestroke approaches top dead center (TDC), i.e., the point at which thepiston 16 is closest to the cylinder head 18, the volume of the mixtureis compressed to one eighth of its initial volume (due to an 8 to 1compression ratio). The mixture is then ignited by an electric sparkfrom spark plug 30.

Referring to FIG. 3, the power stroke follows with both valves 20 and 22still closed. The piston 16 is driven downward (as indicated by arrow44) toward bottom dead center (BDC), due to the expansion of the burnedgas pressing on the crown 24 of the piston 16. Since the spark plug 30is fired when the piston 16 is at or near TDC, i.e. at its firingposition, the combustion pressure (indicated by arrow 56) exerted by theignited gas on the piston 16 is at its maximum at this point. Thispressure 56 is transmitted through the connecting rod 34 and results ina tangential force or torque (as indicated by arrow 58) on thecrankshaft 38.

When the piston 16 is at ifs firing position, there is a significantclearance distance 60 between the top of the cylinder 14 and the crown24 of the piston 16. Typically, the clearance distance is between 0.5 to0.6 inches. For the standard engine 10 illustrated the clearancedistance is substantially 0.571 inches. When the piston 16 is at itsfiring position conditions are optimal for ignition, i.e., optimalfiring conditions. For purposes of comparison, the firing conditions ofthis engine 10 exemplary embodiment are: 1) a 4 inch diameter piston, 2)a clearance volume of 7.181 cubic inches, 3) a pressure before ignitionof approximately 270 pounds per square inch absolute (psia), 4) amaximum combustion pressure after ignition of approximately 1200 psiaand 5) operating at 1400 RPM.

This clearance distance 60 corresponds typically to the 8 to 1compression ratio. Typically, SI engines operate optimally with a fixedcompression ratio within a range of about 6.0 to 8.5, while thecompression ratios of CI engines typically range from about 10 to 16:The piston's 16 firing position is generally at or near TDC, andrepresents the optimum volume and pressure for the fuel-air mixture toignite. If the clearance distance 60 were made smaller, the pressurewould increase rapidly.

Referring to FIG. 4, during the exhaust stroke the ascending piston 16forces the spent products of combustion through the open outlet (orexhaust) valve 22. The cycle then repeats itself. For this prior artfour stoke cycle engine 10, four stokes of each piston 16, i.e. inlet,compression, power and exhaust, and two revolutions of the crankshaft 38are required to complete a cycle, i.e. to provide one power stroke.

Problematically, the overall thermodynamic efficiency of the standardfour stroke engine 10 is only about one third (⅓). That is ⅓ of the workis delivered to the crankshaft, ⅓ is lost in waste heat, and ⅓ is lostout of the exhaust.

As illustrated in FIGS. 3 and 5, one of the primary reasons for this low20 efficiency is the fact that peak torque and peak combustion pressureare inherently locked out of phase. FIG. 3 shows the position of thepiston 16 at the beginning of a power stroke, when the piston 16 is inits firing position at or near TDC. When the spark plug 30 fires, theignited fuel exerts maximum combustion pressure 56 on the piston 16,which is transmitted through the connecting rod 34 to the crankshaftthrow 40 of crankshaft 38. However, in this position, the connecting rod34 and the crankshaft throw 40 are both nearly aligned with thecenterline 52 of the cylinder 14. Therefore, the torque 58 is almostperpendicular to the direction of force 56, and is at its minimum value.The crankshaft 38 must rely on momentum generated from an attachedflywheel (not shown) to rotate it past this position.

Referring to FIG. 5, as the ignited gas expands in the combustionchamber 26, the piston 16 descends and the combustion pressure 56decreases. However, as the crankshaft throw 40 rotates past thecenterline 52 and TDC, the resulting tangential force or torque 58begins to grow. The torque 58 reaches a maximum value when thecrankshaft throw 40 rotates approximately 30 degrees past the centerline52. Rotation beyond that point causes the pressure 56 to fall off somuch that the torque 58 begins to decrease again, until both pressure 56and torque 58 reach a minimum at BDC. Therefore, the point of maximumtorque 58 and the point of maximum combustion pressure 56 are inherentlylocked out of phase by approximately 30 degrees.

Referring to FIG. 6, this concept can be further illustrated. Here, agraph of tangential force or torque versus degrees of rotation from TDCto BDC is shown at 62 for the standard prior art engine 10.Additionally, a graph of combustion pressure versus degrees of rotationfrom TDC to BDC is shown at 64 for engine 10. The calculations for thegraphs 62 and 64 were based on the standard prior art engine 10 having afour inch stroke, a four inch diameter piston, and a maximum combustionpressure at ignition of about 1200 PSIA. As can be seen from the graphs,the point of maximum combustion pressure 66 occurs at approximately 0degrees from TDC and the point of maximum torque 68 occurs approximately30 degrees later when the pressure 64 has been reduced considerably.Both graphs 62 and 64 approach their minimum values at BDC, orsubstantially 180 degrees of rotation past TDC.

An alternative way of increasing the thermal dynamic efficiency of afour stoke cycle engine is to increase the compression ratio of theengine. However, automotive manufactures have found that SI enginestypically operate optimally with a compression ratio within a range ofabout 6.0 to 8.5, while CI engines typically operate best within acompression ratio range of about 10 to 16. This is because as thecompression ratios of SI or CI engines increase substantially beyond theabove ranges, several other problems occur which outweigh the advantagesgained. For example, the engine must be made heavier and bulkier inorder to handle the greater pressures involved. Also problems ofpremature ignition begin to occur, especially in SI engines.

Many rather exotic early engine designs were patented. However, nonewere able to offer greater efficiencies or other significant advantages,which would replace the standard engine 10 exemplified above. Some ofthese early patents included: U.S. Pat. Nos. 848,029; 939,376;1,111,841; 1,248,250; 1,301,141; 1,392,359; 1,856,048; 1,969,815;2,091,410; 2,091,411; 2,091,412; 2,091,413; 2,269,948; 3,895,614;British Patent No. 299,602; British Patent No. 721,025 and ItalianPatent No. 505,576. In particular the U.S. Pat. No. 1,111,841 to Koenigdisclosed a prior art split piston/cylinder design in which an intakeand compression stroke was accomplished in a compression piston12/cylinder 11 combination, and a power and an exhaust stroke wasaccomplished in an engine piston 7/cylinder 8 combination. Each piston 7and 12 reciprocates along a piston cylinder axis which intersected thesingle crankshaft 5 (see FIG. 3 therein). A thermal chamber 24 connectsthe heads of the compression and engine cylinders, with one end beingopen to the engine cylinder and the other end having a valued dischargeport 19 communicating with the compressor cylinder. A water cooled heatexchanger 15 is disposed at the top of the compressor cylinder 11 tocool the air or air/fuel mixture as it is compressed. A set of spacedthermal plates 25 are disposed within the thermal chamber 24 to re-heatthe previously cooled compressed gas as it passes through.

It was thought that the engine would gain efficiency by making it easierto compress the gas by cooling it. Thereafter, the gas was re-heated inthe thermal chamber in order to increase its pressure to a point whereefficient ignition could take place. Upon the exhaust stroke, hotexhaust gases were passed back through the thermal chamber and out of anexhaust port 26 in an effort to re-heat the thermal chamber.

Unfortunately, transfer of gas in all prior art engines of a splitpiston design always requires work, which reduces efficiency.Additionally, the added expansion from the thermal chamber to the enginecylinder of Koenig also reduced compression ratio. The standard engine10 requires no such transfer process and associated additional work.Moreover, the cooling and re-heating of the gas, back and forth throughthe thermal chamber did not provide enough of an advantage to overcomethe losses incurred during the gas transfer process. Therefore, theKoenig patent lost efficiency and compression ratio relative to thestandard engine 10.

For purposes herein, a crankshaft axis is defined as being offset fromthe piston cylinder axis when the crankshaft axis and thepiston-cylinder axis do not intersect. The distance between the extendedcrankshaft axis and the extended piston-cylinder axis taken along a linedrawn perpendicular to the piston cylinder axis is defined as theoffset. Typically, offset pistons are connected to the crankshaft bywell-known connecting rods and crankshaft throws. However, one skilledin the art would recognize that offset pistons may be operativelyconnected to a crankshaft by several other mechanical linkages. Forexample, a first piston may be connected to a first crankshaft and asecond piston may be connected to a second crankshaft, and the twocrankshafts may be operatively connected together through a system ofgears. Alternatively, pivoted lever arms or other mechanical linkagesmay be used in conjunction with, or in lieu of, the connecting rods andcrankshaft throws to operatively connect the offset pistons to thecrankshaft.

Certain technology relating to reciprocating piston internal combustionengines in which the crankshaft axis is offset from, i.e., does notintersect with, the piston-cylinder axes is described in U.S. Pat. Nos.810,347; 2,957,455; 2,974,541; 4,628,876; 4,945,866; and 5,146,884; inJapan patent document 60-256,642; in Soviet Union patent document1551-880-A; and in Japanese Society of Automotive Engineers (JSAE)Convention Proceedings, date 1996, issue 966, pages 129-132. Accordingto descriptions contained in those publications, the various enginegeometries are motivated by various considerations, including power andtorque improvements and friction and vibration reductions. Additionally,in-line, or straight engines in which the crankshaft axis is offset fromthe piston axes were used in early twentieth century racing engines.

However, all of the improvements gained were due to increasing thetorque angles on the power stroke only. Unfortunately, as will bediscussed in greater detail hereinafter, the greater the advantage anoffset was to the power stroke was also accompanied by an associatedincreasing disadvantage to the compression stroke. Therefore, the degreeof offset quickly becomes self limiting, wherein the advantages totorque, power, friction and vibration to the power stroke do not outweigh the disadvantages to the same functions on the compression stroke.Additionally, no advantages were taught or discussed regarding offsetsto optimize the compression stroke.

By way of example, a recent prior art attempt to increase efficiency ina standard engine 10 type design through the use of an offset isdisclosed in U.S. Pat. No. 6,058,901 to Lee. Lee believes that improvedefficiency will result by reducing the frictional forces of the pistonrings on the side walls over the full duration of two revolutions of afour stroke cycle (see Lee, column 4, lines 1016). Lee attempts toaccomplish this by providing an offset cylinder, wherein the timing ofcombustion within each cylinder is controlled to cause maximumcombustion pressure to occur when an imaginary plane that contains botha respective connection axis of a respective connecting rod to therespective piston and a respective connection axis of the connecting rodto a respective throw of the crankshaft is substantially coincident withthe respective cylinder axis along which the piston reciprocates.

However, though the offset is an advantage during the power stroke, itbecomes a disadvantage during the compression stroke. That is, when thepiston travels from bottom dead center to top dead center during thecompression stroke, the offset piston-cylinder axis creates an anglebetween the crankshaft throw and connecting rod that reduces the torqueapplied to the piston. Additionally, the side forces resulting from thepoor torque angles on the compression stroke actually increase wear onthe piston rings. Accordingly, a greater amount of power must beconsumed in order to compress the gas to complete the compression strokeas the offset increases. Therefore, the amount of offset is severelylimited by its own disadvantages on the compression side. Accordingly,large prior art offsets, i.e., offsets in which the crankshaft mustrotate at least 20 degrees past a pistons top dead center positionbefore the piston can reach a firing position, have not been utilized,disclosed or taught. As a result, the relatively large offsets requiredto substantially align peak torque to peak combustion pressure cannot beaccomplished with Lee's invention.

Variable Compression Ratio (VCR) engines are a class of prior art CIengines designed to take advantage of varying the compression ratio onan engine to increase efficiency. One such typical example is disclosedin U.S. Pat. No. 4,955,328 to Sobotowski. Sobotowski describes an enginein which compression ratio is varied by altering the phase relationbetween two pistons operating in cylinders interconnected through atransfer port that lets the gas flow in both directions.

However, altering the phase relation to vary compression ratios imposedesign requirements on the engine that greatly increase its complexityand decrease its utility. For example, each piston of the pair ofpistons must reciprocate through all four strokes of a complete fourstroke cycle, and must be driven by a pair of crankshafts which rotatethrough two full revolutions per four stroke cycle. Additionally, thelinkages between the pair of crankshafts become very complex and heavy.Also the engine is limited by design to CI engines due to the highercompression ratios involved.

Various other relatively recent specialized prior art engines have alsobeen designed in an attempt to increase engine efficiency. One suchengine is described in U.S. Pat. No. 5,546,897 to Bracken entitled“Internal Combustion Engine with Stroke Specialized Cylinders”. InBrackett, the engine is divided into a working section and a compressorsection. The compressor section delivers charged air to the workingsection, which utilizes a scotch yoke or conjugate drive motiontranslator design to enhance efficiency. The specialized engine can bedescribed as a horizontally opposed engine in which a pair of opposedpistons reciprocate in opposing directions within one cylinder block.

However, the compressor is designed essentially as a super charger whichdelivers supercharged gas to the working section. Each piston in theworking section must reciprocate through all four strokes of intake,compression, power and exhaust, as each crankshaft involved mustcomplete two full revolutions per four-stroke cycle. Additionally, thedesign is complex, expensive and limited to very specialized CI engines.

Another specialized prior art design is described in U.S. Pat. No.5,623,894 to Clarke entitled “Dual Compression and Dual ExpansionEngine”. Clarke essentially discloses a specialized two-stroke enginewhere opposing pistons are disposed in a single cylinder to perform apower stroke and a compression stroke. The single cylinder and thecrowns of the opposing pistons define a combustion chamber, which islocated in a reciprocating inner housing. Intake and exhaust of the gasinto and out of the combustion chamber is performed by specializedconical pistons, and the reciprocating inner housing.

However, the engine is a highly specialized two-stroke system in whichthe opposing pistons each perform a compression stroke and a powerstroke in the same cylinder. Additionally, the design is very complexrequiring dual crankshafts, four pistons and a reciprocating innerhousing to complete the single revolution two-stroke cycle. Also, theengine is limited to large CI engine applications.

Accordingly, there is a need for an improved four-stroke internalcombustion engine, which can enhance efficiency by more closely aligningthe torque and force curves generated during a power stroke withoutincreasing compression ratios substantially beyond normally accepteddesign limits.

SUMMARY OF THE INVENTION

The present invention offers advantages and alternatives over the priorart by providing a four-stroke cycle internal combustion engine having apair of pistons in which one piston of the pair is used for the intakeand compression strokes and another piston of the pair is used for thepower and exhaust strokes, with each four stroke cycle being completedin one revolution of the crankshaft.

These and other advantages are accomplished in an exemplary embodimentof the invention by providing an engine with a crankshaft rotating abouta crankshaft axis of the engine. A power piston is slidably receivedwithin a first cylinder and operatively connected to the crankshaft suchthat the power piston reciprocates through a power stroke and an exhauststroke of a four stroke cycle during a single rotation of thecrankshaft. A compression piston is slidably received within a secondcylinder and operatively connected to the crankshaft such that thecompression piston reciprocates through an intake stroke and acompression stroke of the same four stroke cycle during the samerotation of the crankshaft. A gas passage interconnects the first andsecond cylinders. The gas passage includes an inlet valve and an outletvalve defining a pressure chamber therebetween. The outlet valve permitssubstantially one-way flow of compressed gas from the pressure chamberto the first cylinder. Combustion is initiated in the first cylinderbetween 0 degrees and 40 degrees of rotation of the crankshaft after thepower piston has reached its top dead center position.

In an alternative embodiment of the invention the inlet valve and theoutlet valve of the gas passage substantially maintain at least apredetermined firing condition gas pressure in the pressure chamberduring the entire four stroke cycle.

In another alternative embodiment of the invention, the power pistonleads the compression piston by a phase shift angle that issubstantially equal to or greater than 20 degrees and equal to or lessthan 39 degrees.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram of a representative prior art four strokecycle engine, during the intake stoke;

FIG. 2 is a schematic diagram of the prior art engine of FIG. 1 duringthe compression stoke;

FIG. 3 is a schematic diagram of the prior art engine of FIG. 1 duringthe power stoke;

FIG. 4 is a schematic diagram of the prior art engine of FIG. 1 duringthe exhaust stoke;

FIG. 5 is a schematic diagram of the prior art engine of FIG. 1 when thepiston is at the position of maximum torque;

FIG. 6, is a graphical representation of torque and combustion pressureof the prior art engine of FIG. 1;

FIG. 7 is a schematic diagram of an engine in accordance with thepresent invention during the exhaust and intake strokes;

FIG. 8 is a schematic diagram of the engine of FIG. 7 when the firstpiston has just reached top dead center (TDC) at the beginning of apower stroke;

FIG. 9 is a schematic diagram of the engine of FIG. 7 when the firstpiston has reached its firing position;

FIG. 10, is a graphical representation of torque and combustion pressureof the engine of FIG. 7; and

FIG. 11 is a schematic diagram of an alternative embodiment of an enginein accordance with the present invention having unequal throws andpiston diameters.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring to FIG. 7, an exemplary embodiment of a four stroke internalcombustion engine in accordance with the present invention is showngenerally at 100. The engine 100 includes an engine block 102 having afirst cylinder 104 and a second cylinder 106 extending therethrough. Acrankshaft 108 is journaled for rotation about a crankshaft axis 110(extending perpendicular to the plane of the paper).

The engine block 102 is the main structural member of the engine 100 andextends upward from the crankshaft 108 to the junction with the cylinderhead 112. The engine block 102 serves as the structural framework of theengine 100 and typically carries the mounting pad by which the engine issupported in the chassis (not shown). The engine block 102 is generallya casting with appropriate machined surfaces and threaded holes forattaching the cylinder head 112 and other units of the engine 100.

The cylinders 104 and 106 are openings, typically of generally circularcross section, that extend through the upper portion of the engine block102. Cylinders are defined herein as the chambers within which pistonsof an engine reciprocate, and do not have to be generally circular incross section, e.g., they may have a generally elliptical or half moonshape.

The internal walls of cylinders 104 and 106 are bored and polished toform smooth, accurate bearing surfaces sized to receive a first powerpiston 114, and a second compression piston 116 respectively. The powerpiston 114 reciprocates along a first piston-cylinder axis 113, and thecompression piston 116 reciprocates along a second piston-cylinder axis115. The first and second cylinders 104 and 106 are disposed in theengine 100 such that the first and second piston-cylinder axes 113 and115 pass on opposing sides of the crankshaft axis 110 withoutintersecting the crankshaft axis 110.

The pistons 114 and 116 are typically cup shaped cylindrical castings ofsteel or aluminum alloy. The upper closed ends, i.e., tops, of the powerand compression pistons 114 and 116 are the first and second crowns 118and 120 respectively. The outer surfaces of the pistons 114,116 aregenerally machined to fit the cylinder bore closely and are typicallygrooved to receive piston rings (not shown) that seal the gap betweenthe pistons and the cylinder walls.

First and second connecting rods 122 and 124 each include an angle bend121 and 123 respectively. The connecting rods 122 and 124 are pivotallyattached at their top distal ends 126 and 128 to the power andcompression pistons 114 and 116 respectively. The crankshaft 108includes a pair of mechanically offset portions called the first andsecond throws 130 and 132, which are pivotally attached to the bottomopposing distal ends 134 and 136 of the first and second connecting rods122 and 124 respectively. The mechanical linkages of the connecting rods122 and 124 to the pistons 114, 116 and crankshaft throws 130,132 serveto convert the reciprocating motion of the pistons (as indicated bydirectional arrow 138 for the power piston 114, and directional arrow140 for the compression piston 116) to the rotary motion (as indicatedby directional arrow 142) of the crankshaft 108. The first pistoncylinder axis 113 is offset such that it is disposed in the imaginaryhalf plane through which the first crankshaft throw 130 rotates from itstop dead center position to its bottom dead center position. The secondpiston cylinder axis 115 is offset in the opposing imaginary half plane.

Though this embodiment shows the first and second pistons 114 and 116connected directly to crankshaft 108 through connecting rods 122 and 124respectively, it is within the scope of this invention that other meansmay also be employed to operatively connect the pistons 114 and 116 tothe crankshaft 108. For example a second crankshaft may be used tomechanically link the pistons 114 and 116 to the first crankshaft 108.

The cylinder head 112 includes a gas passage 144 interconnecting thefirst and second cylinders 104 and 106. The gas passage includes aninlet check valve 146 disposed in a distal end of the gas passage 144proximate the second cylinder 106. An outlet poppet valve 150 is alsodisposed in an opposing distal end of the gas passage 144 proximate thetop of the first cylinder 104. The inlet check valve 146 and outletpoppet valve 150 define a pressure chamber 148 there between. The inletvalve 146 permits the one way flow of compressed gas from the secondcylinder 106 to the pressure chamber 148. The outlet valve 150 permitsthe one way flow of compressed gas from the pressure chamber 148 to thefirst cylinder 104. Though check and poppet type valves are described asthe inlet and the outlet valves 146 and 150 respectively, any valvedesign appropriate for the application may be used instead, e.g., theinlet valve 146 may also be of the poppet type.

The cylinder head 112 also includes an intake valve 152 of the poppettype disposed over the top of the second cylinder 106, and an exhaustvalve 154 of the poppet type disposed over the top to the first cylinder104. Poppet valves 150,152 and 154 typically have a metal shaft 156 witha disk 158 at one end fitted to block the valve opening. The other endof the shafts 156 of poppet valves 150, 152 and 154 are mechanicallylinked to camshafts 160,162 and 164 respectively. The camshafts 160,162and 164 are typically a round rod with generally oval shaped lobeslocated inside the engine block 102 or in the cylinder head 112.

The camshafts 160,162 and 164 are mechanically connected to thecrankshaft 108, typically through a gear wheel, belt or chain links (notshown). When the crankshaft 108 forces the camshafts 160, 162 and 164 toturn, the lobes on the camshafts 160, 162 and 164 cause the valves150,152 and 154 to open and close at precise moments in the engine'scycle.

The crown 120 of compression piston 116, the walls of second cylinder106 and the cylinder head 112 form a compression chamber 166 for thesecond cylinder 106. The crown 118 of power piston 114, the walls offirst cylinder 104 and the cylinder head 112 form a separate combustionchamber 168 for the first cylinder 104. A spark plug 170 is disposed inthe cylinder head 112 over the first cylinder 104 and is controlled by acontrol device (not shown) which precisely times the ignition of thecompressed air gas mixture in the combustion chamber 168. Though thisembodiment describes a spark ignition (SI) engine, one skilled in theart would recognize that compression ignition (CI) engines are withinthe scope of this invention also.

During operation, the power piston 114 leads the compression piston 116by a phase shift angle 172, defined by the degrees of rotation thecrankshaft 108 must rotate after the power piston 114 has reached itstop dead center position in order for the compression piston 116 toreach its respective top dead center position. Preferably this phaseshift is between 30 to 60 degrees. For this particular preferredembodiment, the phase shift is fixed substantially at 50 degrees.

FIG. 7 illustrates the power piston 114 when it has reached its bottomdead center (BDC) position and has just started ascending (as indicatedby arrow 138) into its exhaust stroke. Compression piston 116 is laggingthe power piston 114 by 50 degrees and is descending (arrow 140) throughits intake stroke. The inlet valve 156 is open to allow an explosivemixture of fuel and air to be drawn into the compression chamber 166.The exhaust valve 154 is also open allowing piston 114 to force spentproducts of combustion out of the combustion chamber 168.

The check valve 146 and poppet valve 150 of the gas passage 144 areclosed to prevent the transfer of ignitable fuel and spent combustionproducts between the two chambers 166 and 168. Additionally during theexhaust and intake strokes, the inlet check valve 146 and outlet poppetvalve 150 seal the pressure chamber 148 to substantially maintain thepressure of any gas trapped therein from the previous compression andpower strokes.

Referring to FIG. 8, the power piston 114 has reached its top deadcenter (TDC) position and is about to descend into its power stroke(indicated by arrow 138), while the compression piston 116 is ascendingthrough its compression stroke (indicated by arrow 140). At this point,inlet check valve 146, outlet valve 150, intake valve 152 and exhaustvalve 154 are all closed.

At TDC piston 114 has a clearance distance 178 between the crown 118 ofthe piston 114 and the top of the cylinder 104. This clearance distance178 is very small by comparison to the clearance distance 60 of standardengine 10 (best seen in FIG. 3). This is because the power stroke inengine 100 follows a low pressure exhaust stroke, while the power strokein standard engine 10 follows a high pressure compression stroke.Therefore, in distinct contrast to the standard engine 10, there islittle penalty to engine 100 to reduce the clearance distance 178 sincethere is no high pressure gas trapped between the crown 118 and the topof the cylinder 114. Moreover, by reducing the clearance distance 178, amore thoroughly flushing of nearly all exhaust products is accomplished.

In order to substantially align the point of maximum torque with maximumcombustion pressure, the crankshaft 108 must be rotated approximately 40degrees past its top dead center position when the power piston 114 isin its optimal firing position. Additionally, similar considerationshold true on the compression piston 116, in order to reduce the amountof torque and power consumed by the crankshaft 108 during a compressionstroke. Both of these considerations require that the offsets on thepiston-cylinder axes be much larger than any previous prior art offsets,i.e., offsets in which the crankshaft must rotate at least 20 degreespast a pistons top dead center position before the piston can reach afiring position. These offsets are in fact so large that a straightconnecting rod linking the pistons 114 and 116 would interfere with thelower distal end of the cylinders 104 and 106 during a stroke.

Accordingly, the bend 121 in connecting rod 122 must be disposedintermediate its distal ends and have a magnitude such that theconnecting rod 122 clears the bottom distal end 174 of cylinder 104while the power piston 114 reciprocates through an entire stroke.Additionally, the bend 123 in connecting rod 124 must be disposedintermediate its distal ends and have a magnitude such that theconnecting rod 124 clears the bottom distal end 176 of cylinder 106while the compression piston 116 reciprocates through an entire stroke.

Referring to FIG. 9, the crankshaft 108 has rotated an additional 40degrees (as indicated by arrow 180) past the TDC position of powerpiston 114 to reach its firing position, and the compression piston 116is just completing its compression stroke. During this 40 degrees ofrotation, the compressed gas within the second cylinder 116 reaches athreshold pressure which forces the check valve 146 to open, while cam162 is timed to also open outlet valve 150. Therefore, as the powerpiston 114 descends and the compression piston 116 ascends, asubstantially equal mass of compressed gas is transferred from thecompression chamber 166 of the second cylinder 106 to the combustionchamber 168 of the first cylinder 104. When the power piston 114 reachesits firing position, check valve 146 and outlet valve 150 close toprevent any further gas transfer through pressure chamber 148.Accordingly, the mass and pressure of the gas within the pressurechamber 148 remain relatively constant before and after the gas transfertakes place. In other words, the gas pressure within the pressurechamber 148 is maintained at least (at or above) a predetermined firingcondition pressure, e.g., approximately 270 psia, for the entire fourstroke cycle.

By the time the power piston 114 has descended to its firing positionfrom TDC, the clearance distance 178 has grown to substantially equalthat of the clearance distance 60 of standard engine 10 (best seen inFIG. 3), i.e., 0.571. Additionally, the firing conditions aresubstantially the same as the firing conditions of the standard engine10, which are generally: 1) a 4 inch diameter piston, 2) a clearancevolume of 7.181 cubic inches, 3) a pressure before ignition ofapproximately 270 pounds per square inch absolute (psia), and 4) amaximum combustion pressure after ignition of approximately 1200 psia.Moreover, the angle of the first throw 130 of crankshaft 108 is in itsmaximum torque position, i.e., approximately 40 degrees past TDC.Therefore, spark plug 170 is timed to fire such that maximum combustionpressure occurs when the power piston 114 substantially reaches itsposition of maximum torque.

During the next 10 degrees of rotation 142 of the crankshaft 108, thecompression piston 116 will pass through to its TDC position andthereafter start another intake stroke to begin the cycle over again.The compression piston 116 also has a very small clearance distance 182relative to the standard engine 10. This is possible because, as the gaspressure in the compression chamber 166 of the second cylinder 106reaches the pressure in the pressure chamber 148, the check valve 146 isforced open to allow gas to flow through. Therefore, very little highpressure gas is trapped at the top of the power piston 116 when itreaches its TDC position.

The compression ratio of engine 100 can be anything within the realm ofSI or CI engines, but for this exemplary embodiment it is substantiallywithin the range of 6 to 8.5. As defined earlier, the compression ratiois the maximum volume of a predetermined mass of an air-fuel mixturebefore a compression stroke, divided by the volume of the mass of theair-fuel mixture at the point of ignition. For the engine 100, thecompression ratio is substantially the ratio of the displacement volumein second cylinder 106 when the compression piston 116 travels from SDCto TDC to the volume in the first cylinder 104 when the power piston 114is at its firing position.

In distinct contrast to the standard engine 10 where the compressionstroke and the power stroke are always performed in sequence by the samepiston, the power stroke is performed by the power piston 114 only, andthe compression stroke is performed by the compression piston 116 only.Therefore, the power piston 116 can be offset to align maximumcombustion pressure with maximum torque applied to the crankshaft 108without incurring penalty for being out of alignment on the compressionstroke. Vice versa, the compression piston 114 can be offset to alignmaximum compression pressure with maximum torque applied from thecrankshaft 108 without incurring penalty for being out of alignment onthe power stroke.

Referring to FIG. 10, this concept can be further illustrated. Here, agraph of tangential force or torque versus degrees of rotation from TDCfor power piston 114 is shown at 184 for the engine 100. Additionally, agraph of combustion pressure versus degrees of rotation from TDC forpower piston 114 is shown at 186 for engine 100. The calculations forthe graphs 184 and 186 were based on the engine 100 having firingconditions substantially equal to that of a standard engine. That is: 1)a 4 inch diameter piston, 2) a clearance volume of 7.181 cubic inches,3) a pressure before ignition of approximately 270 pounds per squareinch absolute (psia), 4) a maximum combustion pressure after ignition ofapproximately 1200 psia and 5) substantially equal revolutions perminute (RPM) of the crankshafts 108 and 38. In distinct contrast withthe graphs of FIG. 6 for the standard prior art engine 10, the point ofmaximum combustion pressure 188 is substantially aligned with the pointof maximum torque 190. This alignment of combustion pressure 186 withtorque 184 results in a significant increase in efficiency.

Moreover, the compression piston's 116 offset can also be optimized tosubstantially align the maximum torque delivered to the compressionpiston 116 from the crankshaft 108 with the maximum compression pressureof the gas. The compression pistons 116 offset reduces the amount ofpower exerted in order to complete a compression stroke and furtherincreases the overall efficiency of engine 100 relative to the standardengine 10. With the combined power and compression piston 114, and 116offsets, the overall theoretical efficiency of engine 100 can beincreased by approximately 20 to 40 percent relative to the standardengine.

Referring to FIG. 11, an alternative embodiment of a split four strokeengine having unequal throws and unequal piston diameters is showngenerally at 200. Because the compression and power strokes areperformed by separate pistons 114, 116, various enhancements can be madeto optimize the efficiency of each stroke without the associatedpenalties incurred when the strokes are performed by a single piston.For example, the compression piston diameter 204 can be made larger thanthe power piston diameter 202 to further increase the efficiency ofcompression. Additionally, the radius 206 of the first throw 130 for thepower piston 114 can be made larger than the radius 208 of the secondthrow 132 for the compression piston 116 to further enhance the totaltorque applied to the crankshaft 108.

While preferred embodiments have been shown and described, variousmodifications and substitutions may be made thereto without departingfrom the spirit and scope of the invention. Accordingly, it is to beunderstood that the present invention has been described by way ofillustration and not limitation.

1. An engine comprising: a crankshaft, rotating about a crankshaft axisof the engine; a power piston slidably received within a first cylinderand operatively connected to the crankshaft such that the power pistonreciprocates through a power stroke and an exhaust stroke of a fourstroke cycle during a single rotation of the crankshaft; a compressionpiston slidably received within a second cylinder and operativelyconnected to the crankshaft such that the compression pistonreciprocates through an intake stroke and a compression stroke of thesame four stroke cycle during the same rotation of the crankshaft; and agas passage interconnecting the first and second cylinders, the gaspassage including an inlet valve and an outlet valve defining a pressurechamber therebetween, the outlet valve permitting substantially one wayflow of compressed gas from the pressure chamber to the first cylinder;wherein combustion is initiated in the first cylinder between 0 degreesand 40 degrees of rotation of the crankshaft after the power piston hasreached its top dead center position.
 2. The engine of claim 1 whereinthe inlet valve and the outlet valve of the gas passage substantiallymaintain at least a predetermined firing condition gas pressure in thepressure chamber during the entire four stroke cycle.
 3. The engine ofclaim 1 wherein the inlet valve permits substantially one way flow ofcompressed gas from the second cylinder to the pressure chamber.
 4. Theengine of claim 1 wherein the power piston leads the compression pistonby a phase shift angle that is substantially greater than 0 degrees. 5.The engine of claim 4 wherein the phase shift angle is substantiallyequal to or greater than 20 degrees.
 6. The engine of claim 4 whereinthe phase shift angle is substantially equal to or less than 39 degrees.7. The engine of claim 4 wherein the phase shift angle is substantiallyequal to or less than 29 degrees.
 8. The engine of claim 5 wherein thephase shift angle is substantially equal to or less than 39 degrees. 9.The engine of claim 5 wherein the phase shift angle is substantiallyequal to or less than 29 degrees.
 10. The engine of claim 1 whereincombustion is initiated in the first cylinder between 5 degrees and 40degrees of rotation of the crankshaft after the power piston has reachedits top dead center position.
 11. An engine comprising: a crankshaft,rotating about a crankshaft axis of the engine; a power piston slidablyreceived within a first cylinder and operatively connected to thecrankshaft such that the power piston reciprocates through a powerstroke and an exhaust stroke of a four stroke cycle during a singlerotation of the crankshaft; a compression piston slidably receivedwithin a second cylinder and operatively connected to the crankshaftsuch that the compression piston reciprocates through an intake strokeand a compression stroke of the same four stroke cycle during the samerotation of the crankshaft; and a gas passage interconnecting the firstand second cylinders, the gas passage including an inlet valve and anoutlet valve defining a pressure chamber therebetween, the inlet valvepermitting substantially one way flow of compressed gas from the secondcylinder to the pressure chamber and the outlet valve permittingsubstantially one way flow of compressed gas from the pressure chamberto the first cylinder, the inlet valve and the outlet valve of the gaspassage substantially maintaining at least a predetermined firingcondition gas pressure in the pressure chamber during the entire fourstroke cycle; wherein combustion is initiated in the first cylinderbetween 5 degrees and 40 degrees of rotation of the crankshaft after thepower piston has reached its top dead center position.
 12. The engine ofclaim 11 wherein the power piston leads the compression piston by aphase shift angle that is substantially greater than 0 degrees.
 13. Theengine of claim 12 wherein the phase shift angle is substantially equalto or greater than 20 degrees.
 14. The engine of claim 12 wherein thephase shift angle is substantially equal to or less than 39 degrees. 15.The engine of claim 12 wherein the phase shift angle is substantiallyequal to or less than 29 degrees.
 16. The engine of claim 13 wherein thephase shift angle is substantially equal to or less than 39 degrees. 17.The engine of claim 13 wherein the phase shift angle is substantiallyequal to or less than 29 degrees.